DE112011103079T5 - Simplified turbocharger with variable geometry with an increased current range - Google Patents

Abstract

A variable geometry turbocharger is simplified but still capable of sustaining the momentum energy. According to a first embodiment, a turbine housing is provided with a pivotal flow control valve which is pivoted about a point near the entrance to the turbine housing. By moving the valve about the pivot point, the effective volume of the turbine housing compressor is changed, effectively reducing the volume of exhaust gas in the compressor, thereby allowing control of the exhaust gas flowing to the turbine wheel. According to the second embodiment of the invention, a rotatable wedge segment within the distributor is rotated from a first position to a second position, whereby the effective volume of the distributor is changed and control of the exhaust gas flowing to the turbine wheel is possible.

Description

FIELD OF THE INVENTION

This invention addresses the need for a turbine flow control device of reduced cost and increased range, and accomplishes this by designing a simplified variable geometry turbocharger housing having a controlled asymmetric flow to the turbine wheel.

BACKGROUND OF THE INVENTION

Turbochargers are one type of forced induction system. They deliver higher density air to the engine intake than would be possible with the normal intake configuration, allowing more fuel to be burned, thus increasing the horsepower of the engine without significantly increasing the weight of the engine. A smaller turbocharged engine can replace a normal naturally aspirated engine with a larger physical size, thereby reducing the mass and aerodynamic frontal area of the vehicle.

Turbocharger ( 1 ) use the exhaust stream ( 100 ) from the engine exhaust manifold, which at the turbine inlet ( 51 ) of a turbine housing ( 2 ) enters the turbine housing to a turbine wheel ( 70 ), which is located in the turbine housing. The turbine wheel is rigidly attached to one end of a shaft. A compressor wheel ( 20 ) is attached to the other end of the shaft and held in place by the clamping load of a compressor nut. The primary function of the turbine wheel is to provide rotational power to drive the compressor. Once the exhaust gas through the turbine wheel ( 70 ) and the turbine wheel has taken energy from the exhaust gas, the spent exhaust gas ( 101 ) through the outlet ( 52 ) from the turbine housing ( 2 ) and is routed to the vehicle exhaust and usually to aftertreatment devices such as catalysts, particulate traps and NO _x traps.

The power developed by the turbine stage is a function of the expansion ratio across the turbine stage, ie, the expansion ratio of the turbine inlet ( 51 ) to the turbine outlet ( 52 ). The range of turbine power is a function of the mass flow through the turbine stage, apart from other parameters.

The compressor stage consists of a wheel and its housing. Filtered air is released by the rotation of the compressor wheel ( 20 ) axially into the inlet ( 11 ) a compressor cover ( 10 sucked). The power generated by the turbine stage and acting on the axle and the wheel drives the compressor wheel ( 20 ) to produce a combination of static pressure and residual kinetic energy and heat. The pressurized gas passes through the compressor outlet ( 12 ) from the compressor cover ( 10 ) and is transferred to the engine intake, usually via an intercooler.

The design of the turbine stage is a compromise between the power required to drive the compressor in different current ranges in the envelope of engine operation, the aerodynamic design of the stage, the inertia of the rotating assembly, of which the turbine is a large part, because the turbine wheel typically made of Inconel, which has a density three times greater than the aluminum of the compressor wheel, the duty cycle of the turbocharger that affects the structural and material aspects of the design, and the near field (exhaust flow) both upstream and downstream of the turbine wheel the blade excitation.

Part of the physical design of the turbine housing is a compressor ( 47 ) or a pair of compressors whose function is to control the inlet conditions for the turbine wheel so that the inlet flow conditions provide the most efficient transmission of power from the energy in the exhaust gas to the turbine wheel developed power, combined with the best transient response characteristics. Theoretically, the exhaust gas flow arriving from the engine is smoothly supplied by the compressor to a vortex wheel centered on the turbine wheel axle. For this purpose, the cross-sectional area of the compressor is ideally at a maximum perpendicular to the current direction and gradually and continuously decreases until it becomes zero. The inner boundary of the compressor can be a perfect circle that is considered the base circle ( 71 ), or in certain cases, such as an in 2A illustrated double compressor ( 48 . 49 ), describe a spiral with a minimum diameter not smaller than 106% of the turbine wheel diameter.

The compressor is characterized by the decreasing radius of the outer boundary of the compressor ( 53 ) and through the inner boundary, as described above, in one on the "XY" axis, as in FIG 4 is defined, and the cross-sectional areas at each station in the plane passing through the "Z" axis, as in FIG 8A is shown defined. The "Z" axis is perpendicular to the plane defined by the "XY" axis and is also the axis of the turbine wheel.

Multiple input compressors can also be created by dividing the compressor area in the circumferential direction. The compressor is driven by axial walls ( 103 . 104 ), which follow the decreasing outer boundary of the compressor, as in 15A is shown.

For the consistency of the product design, a system is used in which the development of the compressor at the intersection "A" ( 4 ) defined as the reference location for the remainder of the compressor. The reference location, section "A" is defined as the intersection at an angle of "P" degrees above the "X" axis of the turbine housing, where the "X" axis, "Y" axis and "Z" Axes specifications of the compressor form are included.

The size and shape of the compressor are defined as follows: The commonly used term A / R represents the ratio between the sectional area in section "A" and the distance from the centroid ( 161 ) of the shaded area of flow to the centerline of the turbocharger 8A sets the position of the centroid ( 161 ) fix the distance R _A to the centerline of the turbocharger. For various elements of a family of turbine housings, the general shape remains the same, but the area in section "A" is different, as is the distance R _A. The A / R ratio is generally used as a "name" for a specific turbine housing to distinguish this turbine housing from others in the same family (with different A / R ratios). In 8A the compressor is that of a typical divided turbine housing which forces the shapes of the compressors to be approximately triangular and of approximately the same area. In the case of a double-flow design (as in 8A shown), the areas in section "A" are the same for both compressors. The centroids ( 160 . 161 ) of the surfaces are at the same radius R _A. The average centroid, ( 163 ) lies on the turbine housing centerline at the same radius R _A because the individual compressors in this section are symmetrical about the bulkhead.

The section "A" is offset by an angle "P" against the "X" axis. The turbine housing is then geometrically decomposed into equal radial sections (often 30 ° to lie at (30x + P °)), and the surfaces (A _AM ) and radii (R _AM ), along with other geometric definitions, in FIG Type of corner radii defined. Using this definition, splines are created at points along the compressor walls, thereby defining the overall shape of the compressor. The wall thickness is added to the shape of the inner compressor, and a turbine housing is defined by this method.

The theoretically optimized compressor shape for a given area is that of a circular cross-section because it has the minimum surface area that minimizes fluid friction losses. However, the compressor does not work alone, but is part of a system so that the requirements of the flow in the levels of in 4 illustrated section "A" to the plane at the intersection "M" and "M" to the tongue affect the operation of the turbine stage. These requirements are often met by other architectural requirements (space availability) outside the turbine housing, the method of locating and mounting the turbine housing to the bearing housing, and the transition from the "A" section to the turbine foot (FIG. 51 ), which in combination force turbine housing compressors to assume a rectangular or triangular cross-section as well as a circular shape or combinations of all shapes. The rectangular shape of the compressor in 1 in which a section "DK" is shown is a result of the request, not just VTG ( 31 ) Turbine blades into the space so that the flow through the turbine blades is optimized and the turbine blades can be moved and controlled by devices outside the turbine housing, but also to minimize the outline of the turbine housing so that the turbocharger fits an engine.

The turbine housing foot is usually of a standard design because it fits exhaust manifolds of many engines. The foot can be placed at any angle to the "compressor" or in any position to the "compressor". The transition from the foot gas passages to the compressor is in a manner that provides the best aerodynamic and mechanical compromise.

The roughly triangular shape of the compressor in 2 in the same sections as those mentioned above, the more typical compressor geometry is for solid turbine housings and wastegate turbine housings. The partition ( 25 ) is added to the aerodynamic "cross-flow" between the compressors in order to try to get a pulse stream from a subdivided distributor ( 36 ) to harvest the pulse energy in the work extracted by the turbine wheel. The pressure pulses in the exhaust manifold are a function of the firing order of the engine.

In commercial practice, turbine housings are typically designed in families (typically 5 to 7 in a family) using turbine wheels of the same diameter or a group of wheels of nearly equal diameter in a given family. You can use the same turbine foot size, although sometimes this feature is specified by the customer. For example, a family of turbine housings for a 63 mm turbine wheel may cover an A / R range of 1.8 to 2.2. 5 shows the area plan for three compressors of a family. The largest compressor is a 1.2 A / R compressor, indicated by the dotted line ( 45 ) is shown. The smallest compressor is a 0.8 A / R compressor, indicated by the broken line ( 46 ) and the middle compressor in the middle of the family is shown by the solid line. The X-axis shows the angle of the cut from 30 ° (section "A") to 360 ° (tongue). The Y-axis shows the cross-sectional area at the respective angle. Typically, there is an 8 to 10% difference in cross-sectional area (in the given case with 12 faces) at the intersection "A" from one A / R to the next A / R in a design family. The compressor outer wall with the largest A / R ( 45 ) out 5 is in 4 as the inner surface of the compressor wall ( 40 ), and the smallest A / R ( 46 ) out 5 is in 4 as area ( 41 ).

Some turbine wheels are specifically designed to curb this pulse energy and convert it to a rotational speed. Accordingly, the conversion of the pressure and velocity from exhaust gas for a pulse turbine impeller in a split turbine casing is greater than converting the pressure and velocity from an equilibrium exhaust gas flow to the turbine wheel speed. The pulse energy is more significant in conventional diesel engines operating at about 2200 RPM with peak torque at 1200 to 1400 RPM than in gasoline engines operating at much higher rotational speeds, often up to 6000 RPM, with peak torque at 4000 RPM, work so that the impulse is not well defined.

The basic turbocharger configuration is that of a fixed turbine housing. In this configuration, the shape and volume of the turbine housing compressor are set at design stage and poured in-situ. Most diesel turbine housings are of the subdivided variety with a radial partition ( 25 ), as in 2 is shown, which separates the two compressors to obtain the pulse energy for the turbine wheel. The length of the partition is typically such that the inner boundary is approximately at the base circle. The closer the tip of the partition wall is to the base circle, the greater the conservation of the pulse energy, but the greater the tendency for the casting in the partition to crack. There are many reasons for this cracking, but the slag that is pushed through the pattern during the casting process is predominant, which means that the integrity of the material near the top of the bulkhead is sub-optimal, and the second reason is the fact that the temperature distribution around the compressors causes the casting to "unwind". The thermal forces that create the "unwinding" of the turbine housing are resisted by the vertical bulkhead, resulting in cracks in the wall. Although a crack causes only minor physical damage, the next step in cracking is that parts of the cast iron bulkhead detach from the casting and be taken up by the turbocharger or engine, which can result in destructive damage.

The next level of higher development after the fixed turbine housing is a wastegate turbine housing. With this configuration, the compressor is cast in place as in the previous fixed configuration. In 2 For example, the turbine housing provided with a wastegate has an opening ( 54 ), which the turbine housing compressor ( 49 ) Fluid-moderately with the turbine housing outlet ( 52 ) connects. Because the opening on the compressor side upstream of the turbine wheel ( 70 ) and the other side of the opening on the outlet side is located downstream of the turbine wheel, the current through the channel connecting these openings bypasses the turbine wheel (FIG. 70 ) so that it does not contribute to the power supplied to the turbine wheel.

The wastegate is in its simplest form a valve ( 55 ), which is a blow-off valve or a pivotable valve similar to the valve in 2 can be. Typically, these valves are actuated by a "simple" actuator which measures the boost pressure or vacuum to activate a diaphragm connected to the valve and operates without specific communication with the engine ECU. The function of the wastegate valve in this manner is to disconnect the tip from the full load boost curve and thereby limit the boost level for the engine. As a result, the effective current to the turbine is substantially reduced, if desired (for example, excessive overdriving of the turbine) Preventing turbine), while the entire range of the turbine housing flow to the turbine wheel is enabled when a full power is desired. The wastegate configuration has no effect on the characteristics of the boost curve until the valve opens. More sophisticated wastegate valves can measure barometric pressure or have electronic override or control, but they also have no effect on the boost curve until they actuate to open or close the valve.

6A shows the gain curve ( 65 ) for a turbine housing with a fixed geometry or a wastegate turbine housing, wherein the wastegate valve has not opened. The X-axis shows the mass flow, and the Y-axis shows the pressure ratio. 6B shows the gain curve ( 67 ) for a wastegate turbine housing with the same A / R as in FIG 6A , wherein the wastegate valve has opened. In 6B it can be seen that the lower mold ( 62 ) of the gain curve ( 67 ) to the point ( 66 ), at which the valve opens, exactly equal to the amplification curve ( 65 ) in 6A is. After this point, the gain curve is flat. While a wastegate may be used to limit the boost levels, its turbine power control characteristics are rudimentary and crude.

An advantageous by-product of the wastegate turbine housing is the ability to reduce the A / R ratio of the turbine housings. Because the upper limit of gain is controlled by the wastegate, reducing the A / R ratio can provide better transient response characteristics while still controlling the upper limit. However, if the wastegate turbocharger has a "simple" actuator that opens only at a pressure or vacuum signal and is actuated at altitude, the critical pressure ratio at which the valve opens will be adversely affected. Because the graph in the actuator measures the boost pressure on one side and the barometric pressure on the other, there is a tendency for the actuator to open later (because the barometric pressure is lower in altitude than at sea level), which is too high Reinforcement for the engine leads. By introducing a turbine housing with a smaller A / R ratio to utilize the wastegate, this A / R reduction also reduces the turbine stage's current range.

Engine boost requirements are the predominant driving elements for selecting the compressor stage. The selection and design of the compressor is a compromise between the engine boost pressure requirement, the mass flow required by the engine, the efficiency demanded by the application, the map width required by the engine and application, the altitude, and the duty cycle to which the engine is subjected is, the cylinder pressure limits of the engine, etc.

The reason that this is important to the operation of a turbocharger is that the addition of a wastegate to the turbine stage allows it to adapt to the low speed range with a smaller turbine wheel and casing. Accordingly, the addition of a wastegate entails the option of reducing inertia. Because reducing the inertia of the rotating assembly typically results in a reduction of particulate matter (PM), wastegates have become commonplace in highway-powered vehicles. The problem is that most wastegates are somewhat binary in their operation, which does not fit well with the linear relationship between engine output and engine speed.

The next higher level of development in turbocharger gain control is VTG (the generic term for variable turbine geometry). Some of these turbochargers have rotating turbine blades, and some have sliding sections or rings. Some designations for these devices are variable turbine geometry (VTG), variable geometry turbine (VGT), variable nozzle turbine (VNT) or single variable geometry (VG).

VTG turbochargers use adjustable guide turbine blades ( 3A and 3B ) rotatably connected to a pair of vane rings and / or the nozzle wall. These turbine blades are adjusted to control the exhaust back pressure and turbocharger speed by modulating the exhaust flow to the turbine wheel. In 3A are the turbine blades ( 31 ) in the minimum open position. In 3B are the turbine blades ( 31 ) in the maximum open position. The turbine blades may be rotationally driven by engaging arms engaging a union ring that may be located above the upper turbine blade ring. For the sake of clarity, these details have been omitted from the drawing. VTG turbochargers have a large number of very expensive alloy components which must be assembled and placed in the turbine housing so that the guide turbine blades remain properly positioned with respect to the exhaust flow channel and turbine wheel over the range of thermal operating conditions to which they are exposed. The temperature and Corrosion conditions force the use of exotic alloys in all internal components. These materials are very expensive to procure, process and weld (if necessary). Because the VTG design can change turbocharger speed very quickly, expensive software and controls are needed to prevent unwanted outliers. This results in costly actuators. While VTGs of various types and configurations have been widely used to control both turbocharger boost levels and turbine surge pressure levels, the cost of hardware and implementation is high.

The cost of a typical VTG turbocharger in the same production volume is 270% to 300% of the cost of solid geometry turbochargers of the same size. This disparity is due to a number of relevant factors, from the number of components, the materials of the components, the accuracy required in the manufacture and machining of the components, to the speed, accuracy and repeatability of the actuator. The diagram in 7 shows the cost in comparison for the range of turbochargers from solid turbochargers to VTG turbochargers. A column "A" represents the comparison cost of a solid turbocharger for a given application. A column "B" represents the cost of a wastegate turbocharger for the same application, and a column "C" represents the cost of a conventional VTG turbocharger the same application.

It can therefore be seen that there is a need for a relatively inexpensive turbine flow control device that fits in terms of cost between waste gates and existing VTGs, both for technical reasons and in terms of cost drivers.

SUMMARY OF THE INVENTION

The present invention relates to a simplified, low cost, variable geometry turbocharger and, more particularly, to a turbine flow control device which utilizes a split turbine housing with asymmetric compressor A / R in conjunction with a power modulating device to change the effective exhaust mass flow to the turbine wheel as the turbine stage current range increases becomes. By controlling the exhaust mass flow that the turbine housing directs to the turbine wheel with a set of asymmetrically configured compressor throat areas and controlling the flow through the two compressors with a relatively simple flow control device, the flow range may be in a range that provides the symmetrically configured compressor throat area without the flow control device can be exceeded, both widened and controlled.

The variable geometry turbocharger is simplified but still able to sustain the momentum energy. According to a first embodiment, a turbine housing is provided with a pivotal flow control valve which pivots about a point near the entrance to the turbine housing. Moving the valve about the pivot point progressively blocks the flow through the turbine housing at the large compressor, thereby encouraging flow to the small compressor and continuing from there to the turbine wheel, thereby causing the turbine housing to effectively lose the larger compressor acts as a turbine housing with a smaller A / R ratio. According to the second embodiment of the invention, a rotatable butterfly shaped flow control valve pivots within the compressor which pivots about the center of the valve plate and about the centerline to change and thereby cause the flow from the large compressor to the small compressor and on to the turbine wheel in that the turbine housing acts as a turbine housing with a smaller A / R ratio.

Tests conducted by the inventors have found that a 60/40 split of the "A" section surfaces with the hub side at 60% and the air duct side at 40% produced a desirable mass flow split with the restrictor fully open. The asymmetric turbine housing has a larger left or hub side compressor ( 48 ) and a smaller or air channel side compressor ( 49 ) axially around a partition wall ( 25 ) are arranged.

BRIEF DESCRIPTION OF THE DRAWING

The present invention is illustrated by way of example and not limitation in the accompanying drawings in which like reference numerals refer to like parts. Show it:

1 a cross section of a typical VTG turbocharger,

2 a pair of cross sections of a typical wastegate turbocharger,

the 3A , B is a pair of cross sections of a typical VTG turbocharger,

4 a cross section of a typical solid turbine housing, wherein radial construction lines are shown,

5 a diagram of a cross-sectional area development,

the 6A , B the compressor maps for a typical fixed turbocharger and a wastegate turbocharger,

7 a diagram showing the relative cost of turbochargers

the 8A , B are the cross-sections of two types of compressor at section "A",

9 a view of an asymmetric turbine housing to a manifold,

the 10A , B are two sectional views of the restriction device on a circumferentially divided housing,

the 11A B shows two views of a variation of a restriction device on a circumferentially divided housing,

12 a representation of cracking in a turbine housing,

13 a sectional view of closed slots in a turbine housing partition,

14 a sectional view of open slots in a turbine housing partition,

the 15A , B show two views of the third embodiment of a radially divided housing,

16 a diagram showing a mass flow

17 the veil areas for a sample of production turbine wheel diameters and

18 the relationship between the cross-flow area and D ₃ for different turbine stages.

DETAILED DESCRIPTION OF THE INVENTION

As discussed above, variable geometry mechanisms tend to more than double the cost of the simple turbocharger. The inventors sought to be able to more modulate the exhaust gas flow to the turbine wheel more cost-effectively. Therefore, the inventors have experimented with designs having subdivided compressor regions combined with a drag device to provide both a cost effective and technically effective alternative for controlling the required wide range of exhaust gas flow to the turbine. In addition to the above advantages, the inventors sought to provide a low flow range adapted turbocharger that would provide optimized transient response of the turbocharger (and thus the engine) for low currents while still capable of operating at other than low flow conditions cost-effective turbocharger to deliver the high currents required by the engine. This goal keeps the gas velocities in the ideal range where the step efficiencies are maximized.

When a turbocharger is adapted to the maximum power requirement of an engine, the power requirements are met over the entire engine operating range. The problem is that matching the turbocharger to the maximum power requirement means that the size of the turbine housing compressor (and thus the current) is too high for low-current areas of the engine. The transient response characteristics of the turbocharger are sluggish because the entire compressor must be filled to supply power to the turbine wheel. Because reducing the A / R ratio of a turbocharger turbine housing to meet the low-current requirement would mean that the typical speed-limiting turbocharger is unable to handle the high-end high end demand To provide sufficient power to the engine operating areas, the inventors have realized that it is necessary to provide a novel variable geometry turbocharger. Further, in today's EGR (Exhaust Gas Recirculation) requirements, OEM manufacturers use large amounts of EGR at part load (eg, 40% load) and no EGR at high speed, but still want the best possible performance from a market standpoint Deliver full load. High EGR at low speed or at partial load requires a low mass flow. The best possible performance at the nominal point without EGR requires a high mass flow, so it can be seen that the mass flow range of the turbine must be able to meet the current requirements at these two extremes.

The shapes and dimensions of the turbine housing compressor are defined by the area of section "A", and all features and dimensions downstream of section "A" are controlled by the features and dimensions in section "A". This system is used for the consistency of the design in turbochargers designed and manufactured by a turbocharger manufacturer.

In accordance with the present invention, the inventors provide a novel turbine design that is capable of producing a wider turbine flow area than would be available with equal area compressors.

By controlling the mass flow through the turbine housing, the inventors sought to control the mass flow of gas flowing through the turbine housing to the turbine wheel. When the engine operates under a low-speed, low-load condition, the boosting level required to supply the required combustion gas (air) is relatively low. When the engine is under a high-speed, high-load condition, the boost level required to power the engine under these load conditions is high. As the engine transitions from low load conditions to high load conditions, the turbocharger must supply an increasing volume of air at an increased pressure ratio. Because the compressor stage is driven by the turbine stage, the exhaust mass flow required to meet the requirements of the engine (and hence the compressor) must change. That is, under the condition of a low load and a low speed of the engine, the engine exhaust output with respect to the mass flow is low. Under the condition of high load and high engine speed, the engine output is high in terms of mass flow. In the transition stage, the exhaust gas mass flow must change from low to high.

The problem is that the turbine stage must be adapted to both of the basic engine conditions described above, in addition to the EGR requirements, to allow the turbocharger to provide the required flow and pressure ratio under all of these conditions. To force the turbocharger to change speed quickly, experts would choose a turbocharger with a turbine housing with a small A / R ratio. To provide the required flow and pressure ratio under high load and high speed conditions, a turbocharger with a turbine housing having a larger A / R ratio would be selected. The former turbine housing with a small A / R ratio provides good transient response characteristics, but results in insufficient mass flow to the turbine stage to meet the requirement of a high speed, high load compressor. The latter high A / R ratio turbine housing meets the mass flow requirement for the turbine stage for the requirement of high speed, high load boost, but is unable to accelerate the turbine wheel fast enough to be acceptable generate transient response.

Obviously, it would be nice to have a system with two turbochargers, one larger and one smaller, and able to switch between the two. However, such a system would be costly, it would create a large "heat sink", occupy much space in the engine compartment and increase the mass of the vehicle.

A suitably adjusted turbine stage with a small A / R ratio acting alone provides an acceptable transient response, albeit at the cost of a higher back pressure than a turbine stage adapted to the high load and high speed condition. In a non-EGR engine, the presence of a high back pressure is detrimental to the pressure differential across the engine and hence to the efficiency of the engine. In a high pressure loop EGR engine configuration (versus a low pressure EGR engine configuration), the high back pressure in the exhaust system is part of the solution to remove the exhaust gas from the exhaust side of the engine to the intake side of the engine at one Pressure amplification occurs to drive. A large turbine case A / R ratio for a given set of engine parameters develops a lower exhaust back pressure than would be the case with a turbine case having a smaller A / R ratio under the same set of engine parameters. Therefore, the ability to change the effective A / R ratio of the turbine housing allows a single turbocharger to meet both the current and dynamic pressure requirements of a low speed and low load condition and a high speed, high load condition.

By controlling the exhaust mass flow that the turbine housing directs to the turbine wheel with a set of asymmetrically designed compressor cross-sectional areas and controlling the flow through the two compressors with a relatively simple power control device, the current range can be both widened and controlled in a manner that is symmetrically designed Compressor cross-section areas without the flow control device available range is exceeded.

After initially experimenting with a symmetrically sectioned compressor turbine housing, the inventors next experimented with asymmetrically divided turbine housings and found that by replacing one of the compressors with another compressor having a smaller A / R ratio, the current range falls and the maximum current range therethrough Compressor also drops. Similarly, the maximum current range of this compressor would increase by replacing one compressor with another compressor having a larger A / R ratio. By forming a combination of a larger and a smaller compressor and controlling the degree of blockage of the larger compressor, the flow range of the turbine housing according to the invention exceeds that of the original prototype turbine housing with a symmetrically divided compressor. In 16 represent the bars ( 22 ) with the horizontal hatching the mass flow of a turbine housing with compressors of the same area (50-50) and the bars ( 23 ) with the vertical hatches the mass flow of a turbine housing with compressors asymmetric surfaces (60-40). Although the mass flows are the same when the restriction valve is fully open, the mass flow of the asymmetric A / R configuration in the fully closed position of the restriction valve (ie, the position where the flow into the larger compressor is effectively blocked) is smaller than the mass flow of the corresponding one A / R configuration. The sum of areas in section "A" is within 0% to 3% for both configurations, while the change in mass flow is in the range of 10% to 13%.

8A shows a typical symmetrical turbine housing compressor configuration in which the centroids ( 160 . 161 ) of the two compressors are in the same radius R _A from the center line. Because the turbine housing is symmetrical, the effective centroid ( 163 ) of both compressors in the partition wall between the compressors. 8B FIG. 14 shows an example in which the left compressor has an A / R size, that of the symmetrical turbine housing 8A exceeds, and in which the right-hand compressor has a smaller A / R size in the area, namely by 2 A / R sizes is smaller than the left compressor. In this case, the centroid of the right compressor is at a radius R _C from the centerline and axially closer to the centerline of the turbine housing. The centroid of the left compressor is at a radius R _B from the centerline and axially farther from the centerline of the turbine housing. The effective centroid ( 164 ) of both compressors in the turbine housing is now offset by a radius R _D to the left of the center line of the partition wall.

In order to produce an optimum asymmetric turbine casing, the inventors have several options of compressor sizes from a compressor A / R ratio up or a compressor down from compressors of the same size or from compressors of the same size, the hub side being one compressor-A / R ratio increased and the air duct side was reduced by a compressor A / R ratio considered. Tests conducted by the inventors found that the latter solution, which was a 60/40 split of "A" sectional areas with the hub side at 60% and the air duct side at 40%, produced the desired mass flow ratio with the restrictor fully open.

In all subdivided turbine housings, there exists a "cross-flow veil" between the tip of the bulkhead, at the minimum diameter, and the tips of the turbine wheel. In order to minimize the turbine wheel excitation caused by the action of the rotating turbine blades which cause the static tongue ( 26 ) in 4 As a rule of thumb, the depth of the bulkhead typically does not extend closer to the tips of the turbine blades than a ratio of 120% to 150% of the turbine tip diameter D ₃ . This ratio D _bc / D ₃ is usually determined by in-house design rules and technical objectives. The diameter D _bc is known as the base circle. Because the tips of the turbine casing divider tend to crack when the molten cast iron slag is forced into the tip, the innermost or minimum diameter of the divider wall is typically not less than 120% to 150% of the turbine tip diameter. This "veil" between the partition and the turbine wheel allows crossflows of exhaust gas between the two compressors as well as crosstalk between the pulses in the exhaust stream, the latter being the reason why the partition is provided.

For a turbine stage with a base circle whose diameter is 120% of the turbine wheel diameter, there exists a "crossflow veil" with an area that is 70% to 105% of the area of both compressors in a symmetric configuration, in section "A" for a turbine housing family of FIG A / R. For a turbine stage with a base circle whose diameter is 150% of the turbine wheel diameter, the "crossflow veil" has an area that is 199% to 299% of the area of both compressors in a symmetric configuration, in section "A" for a turbine housing family of 5A / R. From this analysis, it can be seen that the veil surface can provide a very large cross-sectional area for cross-flow from one compressor to another.

Because the haze area is a function of both the turbine wheel diameter D ₃ and the minimum position of the bulkhead D _bc , the haze area varies for different values of D ₃ . 17 shows the fog areas for a sample of production turbine wheel diameters of 64 mm to 96 mm. The veil surfaces ( 133 ) are separated by an upper boundary line ( 131 ) and a lower boundary line ( 132 ) included or limited by them. As expected, the area of the haze area increases as D ₃ increases. In 17 are the turbine wheel diameter D ₃ as ( 123 ), and a line ( 124 ) shows the trend of the turbine wheel diameter D ₃ for the analyzed turbos. This diagram also contains ratios of D _bc / D ₃ ranging from 1.25 to 1.35.

The inventors have found, by testing with a 64 mm turbine wheel, that for an asymmetrically designed 60/40 compressor combination with a restriction valve, the optimal cross flow area includes the openings in the dividing wall plus the area of the "basement curtain" (by the difference between the area under D _bc and the area under D ₃ determines) a surface with a ratio of 289.6% of a single symmetrical compressor cross-sectional area in the section "a" (ie, the half of the area in section "a") was. This is to be compared with the typical cross-flow area of the same sized turbine housing without slots or openings with the same ratio of D _bc / D ₃ , which has a cross flow area of only 182.6% of the area half in section "A".

As in the case of the relationship between the haze surfaces and D ₃ and D _bc , the total cross-flow area (FIG. 122 ) is influenced not only by D ₃ and D _bc but also by the variation of the area of a single compressor in section "A". The cross-flow surfaces ( 122 ) are separated by an upper boundary line ( 126 ) and a lower boundary line ( 127 ) limited. The diagram in 18 shows the relationship between the cross-flow area and D ₃ ( 123 ) for various turbine stages analyzed by the inventors.

To select a cross flow area, determine the value of D ₃ , the diameter of the turbine wheel in inches. The example is that of a 76 mm (2,992 '') measuring turbine wheel, which is considered a horizontal line ( 128 ) is shown. From the turbine wheel diameter, the vertical line intersects ( 129 ), which determines the turbine wheel diameter ( 123 ), the lower boundary line ( 127 ) and the upper boundary line ( 128 ). The cross flow area is considered the vertical segment ( 130 ) of the vertical line ( 129 ) between the lower and the upper boundary line ( 127 and 126 ).

The formula that generates the data points that are in the in the 17 and 18 plotted diagrams could be the following:

As in 9 is shown in the low-current condition, a pivotable valve element ( 72 ) to limit the current to the larger compressor ( 48 ) on the side of the hub or the bearing housing, which the flow from the distributor through the smaller compressor ( 49 ) on the side of the air duct or the outlet channel to the turbine wheel ( 70 ) urges.

The asymmetric turbine housing has a larger left or hub side compressor ( 48 ) and a smaller right or air channel side compressor ( 49 ) axially around a partition wall ( 25 ) are arranged. A current limiter, in this case a swiveling valve element ( 72 ), is in the connecting surfaces of the foot ( 37 ) of the central portion of the distributor and the turbine housing foot ( 51 ). Although the inventors have chosen this configuration for cost reasons and for technical reasons, the restrictor could also be used in the hub exhaust manifold ( 34 ) are located.

As in 10B is shown, is located under the high current condition, the pivoting valve element ( 72 ) in the middle position, wherein neither the larger compressor ( 48 ) on the side of the hub even the smaller compressor ( 49 ) on the side of the air duct is preferred to allow the maximum flow to the turbine wheel. Under this high-current condition, the pivotable valve element ( 72 ) of the current-limiting device with the partition wall ( 25 ) of the turbine housing downstream of the turbine base ( 51 ). In the position of minimum current (indicated by the broken line in 10B shown) is the pivotable valve element ( 72 ), preferably by a on the actuator arm ( 73 ) applied force to the axis ( 74 . 78 ) of the device to the closed position, so that it is the exhaust gas flow to the large compressor ( 48 ) and causes the exhaust gas flow through the small compressor ( 49 ) flows. The current limiter can be modulated to any position between the fully open and fully closed positions.

A sectional view of this version of the current limiter device is shown in FIG 10A shown. In this view, it can be seen that the current limiter plate according to the preferred embodiment of the invention with two cylindrical bearing surfaces for pivoting and an actuator arm ( 73 ) is made for position control. One side of the in the connection of the turbine housing foot ( 51 ) and the exhaust manifold foot ( 37 ), wherein a bearing surface is received, is a blind bore ( 77 ), while the other ( 75 ) is an open hole. On the side of the open bore provides a piston ring ( 76 ) Not only an axial orientation for the current limiting device, but also a gas seal. The arrangement of the actuator arm ( 73 ) can be optimized to meet architectural constraints.

The inventors have found that the ratio between boost pressure and back pressure and back pressure alone increased as a function of engine speed and engine load, both at sea level and altitude, thereby making the current limiter device in the exhaust system an ideal control parameter. When the swingable current limiter is rotated toward the closed position, the turbine housing acts as if it were a turbine housing with a smaller A / R ratio than in a case where the current limiter is in the open position. This causes the exhaust back pressure to increase, which is required for an EGR flow from the exhaust side of the engine to the intake side of the engine. Accordingly, the rotation of the current limiter may be used to develop a pressure difference (from the exhaust side of the engine to the intake side of the engine) to assist the EGR flow from the exhaust side of the engine to the intake side of the engine.

According to the first embodiment of the invention, the effective mass flow to the turbine wheel is controlled by a current limiter, which is pivoted about a point in the turbine housing inlet or foot, so that the pivotable valve element (FIG. 72 ) of the current limiter in the open position in a line with the partition wall ( 25 ) of the turbine housing, thereby minimizing the restriction on exhaust flow. If a greater restriction or smaller mass flow to the turbine wheel is required, the swing arm ( 73 ) so that it moves around its axis ( 74 . 78 ), thereby causing the pivotable valve element ( 72 ) the flow of exhaust gas to the large compressor ( 48 ), causing a modulatable reduction in mass flow to the turbine wheel.

At one in the 11A and 11B In the illustrated variation of the first embodiment of the invention, the current limiter takes the form of a butterfly valve ( 80 ), the moment on the swivel arm ( 81 ), which allows the possible use of a lower force and thus a lower cost actuator. In 11A 1 which is a sectional view of the first variation of the first embodiment of the invention, the configurations of the bearing surfaces and the piston ring are the same as in the first embodiment. In 11B In the case of the butterfly configuration, the swivel location is approximately in the middle of the flow path to the large or hub-side compressor ( 48 ), so that the rotation of the Butterfly valve ( 77 ) about its axis ( 74 . 84 ) an adjustable current limit to the hub side compressor ( 48 ), whereby the flow to the air channel side compressor ( 49 ) is favored. In the case of this variation of the first embodiment of the invention, the butterfly valve is in the 11B aligned position of a restriction to a minimum flow with the flow through the compressor, so that the tips of the butterfly valve the hub-side compressor ( 48 ) or shade. A butterfly valve solution has the advantage of lower actuator loads because the moments on either side of the pivot point cancel each other out.

When the current limiter is in the partially open position, the flow from the air duct side (smaller) compressor ( 49 ) to the hub-side (larger) compressor ( 48 ) be further facilitated by either the length of the partition ( 25 ) or by forming slots in the partition wall.

Typically, the turbine housing partitions in the world of commercial diesel engines, where the product can be expected to travel a million miles, are prone to cracking. The inventors have seen an opportunity to mechanically minimize this susceptibility to cracking in the bulkhead by incorporating precast stress relieving features in the bulkhead. 12B shows one along the section AA 12A considered turbine housing. This segmental subdivision is typically performed to assess the condition of the turbine housing after a thermal cycle qualification test wherein the turbocharger is subjected to extreme temperature cycling in an effort to determine its crack resistance. In 12B are the cracks ( 87 ) typical of a commercial diesel turbine housing in the partition wall area.

The inventors have suggested that if "voltage reducers" in the form of slots or openings are poured into the dividing wall, these openings would minimize not only the tendency to crack, but also a current path from the unmodulated air channel side compressor to partial closing of the restricting valve would provide modulated hub-side compressor. This additional power path provides power to the turbine wheel over a larger circumferential distance or area than would be possible without the slots or openings.

According to the in 13 illustrated second embodiment of the invention, the effective mass flow to the turbine by a current limiter in an asymmetric turbine housing with formed in the partition cross-flow openings ( 88 ) controlled. According to the preferred second embodiment of the invention, the area of the openings through the radial ( 89 ) of the leading edge, the radial ( 90 ) of the trailing edge, the circular segment ( 92 ) of the inner edge and the spiral ( 91 ) limits the outer edge for each opening. The sum of the areas of the transverse flow openings in the turbine housing is approximately equal to the area of the modulated compressor in section "A". Important is the sum of the areas of the openings and not the geometry of the openings.

At an in 14 In the illustrated variation of the second embodiment of the invention, the effective mass flow to the turbine wheel is determined by a current limiter in an asymmetric turbine housing with cross-flow slots formed in the bulkhead (US Pat. 95 ) controlled. According to the preferred second embodiment of the invention, the area of the slots through the radial ( 98 ) of the leading edge, the radial ( 99 ) of the trailing edge and the spiral ( 91 ) bounded the outer edge for each opening and is the inner boundary by the base circle ( 71 ). The sum of the areas of the cross-flow slots in the turbine housing is approximately equal to the area of the modulated compressor in section "A". Important is the sum of the areas of the slots and the cross-flow area within the partition wall tip, not the geometry of the slots. According to the preferred second embodiment of the invention, the outer boundary ( 97 ) of the slot may be characterized by a keyhole configuration as the outer termination of the slot to minimize the tendency of the slot to act as a tensioner and initiate tearing.

Multi-flow turbine casing with the compressor dividing wall parallel to the turbocharger axis, d. H. Axial surfaces instead of radial surfaces as in the simple double-flow turbine housing, are not uncommon. The inventors have seen the opportunity to use similar logic for multi-flow turbine housings with asymmetric compressor surfaces in conjunction with a current limiter to further cost-effectively broaden the current range of a turbine stage with this type of turbine housing.

According to the third embodiment of the invention, a three-flow turbine casing is preferably used, as it is in 15A is shown. Two axial compressor partitions ( 103 . 104 ) are formed in the turbine housing so that the ratio of flows through the unrestricted adjacent compressors from outside to inside ranges from about 70% to 20% to 10%. These proportions can be changed depending on the requirements. The ratio of the flows is important only in that the sum of the open areas of the modulated compressors is equal to the area of the modulating restrictor valve. There is a current limiting valve. The flow restrictor pivots about a point in the turbine housing inlet or foot so that in the open position, the plate (Fig. 89 ) of the current limiter with the wall of the outer compressor of the turbine housing, whereby the restriction of the exhaust gas flow is minimized. If a greater restriction or smaller mass flow to the turbine wheel is required, the swing arm ( 73 ) so that it rotates about its axis, thereby causing the plate ( 89 ) obstructs the flow of exhaust gas first to the outer compressor (s) followed by the middle compressor. In this way, the effective mass flow to the turbine wheel is controlled by a current limiter, which allows a modulatable reduction in mass flow to the turbine wheel.

In a variation of the third embodiment of the invention, the partitions ( 106 . 107 ) with slots ( 108 ) to allow a flow from the outer compressors to the internal compressors and then the turbine wheel ( 70 ) reached. The slots ( 108 ) also allow mass flow modulation, but with a more consistent and cheaper power distribution to the turbine wheel.

Claims (14)

A variable flow turbocharger comprising: (a) a turbine housing ( 2 ) with a turbine inlet section on a turbine housing foot ( 51 ) for coupling to an exhaust manifold to receive an exhaust gas stream, a turbine outlet section (US Pat. 52 ) and a compression chamber between the turbine housing foot and the outlet section, (b) a turbine runner ( 70 ) having a plurality of blades located within the turbine housing for receiving an exhaust gas flow from the compression chamber; (c) at least one compressor separation wall for moving the compression chamber into a larger compression chamber (10); 48 ) and a smaller compression chamber ( 49 and (d) exhaust flow control valve means arranged to control the degree of blockage of the exhaust gas flowing into the larger compression chamber. The variable flow turbocharger of claim 1, wherein the exhaust flow control valve means are variably adjustable. The variable flow turbocharger of claim 1, wherein the compressor dividing wall has openings through which the exhaust flow can pass between the larger and smaller compressor chambers. A variable flow turbocharger according to claim 1, wherein the partition wall is an axial partition wall. A variable flow turbocharger according to claim 1, wherein the partition wall is a radial partition wall. The variable flow turbocharger of claim 1, wherein the volume of the larger compressor is at least about 55% of the divided compressor space and the volume of the smaller compressor is at most about 45% of the divided compressor space. The variable flow turbocharger of claim 1, wherein the volume of the larger compressor is about 55 to 65%, and the volume of the smaller compressor is about 45 to 35% of the divided compressor. A variable current turbocharger according to claim 1 having at least two axial or radial partitions. A variable flow turbocharger according to claim 1, wherein the partition wall is radial, and wherein the smaller compressor ( 49 ) on the outlet side and the larger compressor ( 48 ) is opposite to the outlet side. The variable flow turbocharger of claim 1, wherein the exhaust flow control valve means comprises a pivotable valve member (10). 72 ), which is adapted between an open position, in the the power to the larger compressor ( 48 ) and a blocking position in which the flow to the larger compressor ( 48 ) is substantially blocked, to pivot. An internal combustion engine having an exhaust manifold and a variable flow turbocharger in fluid communication with the exhaust manifold, the variable capacity turbocharger comprising: (a) a turbine housing ( 2 ) with a turbine inlet section, a turbine outlet section ( 52 (b) a turbine runner located in the turbine housing and having a plurality of blades; (c) at least one compressor separation wall extending to the vicinity of the turbine inlet portion, into a larger compression chamber (US Pat. 48 ) and a smaller compression chamber ( 49 ) (d) a subdivided exhaust manifold ( 36 with a first exhaust gas introduction passage for introducing exhaust gases into the larger compression chamber and a second exhaust gas introduction passage for introducing exhaust gases into the smaller compression chamber and (e) exhaust gas flow control valve means located at least in the first exhaust gas introduction passage, by the degree of blockage of the gases flowing into the first exhaust gas introduction passage Control exhaust gases. The variable flow turbocharger of claim 11, wherein the exhaust flow control valve means are in the turbine housing. The variable flow turbocharger of claim 11, wherein the exhaust flow control valve means are in the exhaust manifold coupled to the turbine housing. The variable flow turbocharger of claim 11, wherein the exhaust flow control valve means are located between the exhaust manifold and the turbine housing.

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